According to the prior art, automatic transmissions, particularly for motor vehicles, comprise planetary gear sets that are shifted using friction elements or shift elements such as clutches and brakes, and typically are connected to a start-up element, such as a hydrodynamic torque converter or a fluid coupling, that is subject to a slip effect and is provided optionally with a lock-up clutch.
Automatically shiftable vehicle transmissions of planetary design are already generally described numerous times in the prior art and are continually undergoing further development and improvement. These transmissions should have a relatively simple design, in particular requiring a low number of shift elements, and minimize the need for double shifting when sequential shifting is performed, that is, avoiding engaging or disengaging two shift elements, thereby ensuring that only one shift element is ever switched when shifting is performed in defined groups of gears.
The document DE 2721719 A1 describes a multi-stage transmission in planetary design having six forward gears and one reverse gear comprising three minus planetary gear sets, called the first, second and third planetary gear sets in the following, disposed in a housing, six rotatable shafts, called drive shaft, output shaft, third, fourth, fifth and sixth shafts in the following, and five shift elements. Here, the sun gear of the first planetary gear set is connected to the drive shaft, which, via a first clutch, can be releasably connected to the sixth shaft connected to the sun gear of the second planetary gear set and to the sun gear of the third planetary gear set, and via a second clutch, can be releasably connected to the fifth shaft, connected to the carrier of the second planetary gear set and the ring gear of the third planetary gear set, and which can be coupled via a third brake to the housing. With the known transmission it is also provided that the carrier of the first planetary gear set is connected to the fourth shaft, which is connected to the ring gear of the second planetary gear set and can be coupled via a second brake to the housing, and that the ring gear of the first planetary gear set is connected to the third shaft, which can be coupled via a first brake to the housing, wherein the output shaft is connected to the carrier of the third planetary gear set. The brakes and clutches of the transmission are implemented as friction engaged shift elements, particularly as multi-disk shift elements.
Because two engaged shift elements are required for each gear with the transmission according to the document DE 2721719 A1, with each gear three friction engaged shift elements are disengaged, which disadvantageously results in undesired drag torques that negatively influence the efficiency of the transmission.
Further, it is provided that the first clutch is engaged for the first four forward gears, wherein the third brake is required only for implementing the first forward gear and is shifted into the power flow. This means that the first clutch and the third brake, because they are required for the first forward gear, are designed such that they support the entire engine torque including the maximum conversion. For the further gears of the transmission, a substantially smaller design of the shift elements would be sufficient.
Engaging a gear in the first forward gear, the coasting and tractive downshift from second gear into the first forward gear, and the coasting and tractive downshift from fifth gear in to the fourth forward gear, are qualitatively negatively influenced due to the maximum design of the first clutch and the third brake. In order to optimize the shift quality with these shifts, the number of disks of the first clutch and the third brake is reduced according to the prior art, which, however, disadvantageously results in reduced transfer capability of these shift elements in the first forward gear.
From the prior art, for example from the documents DE 10 2008 000 429 A1 and DE 10 2007 022 776 A1 from the applicant, transmissions in planetary design are known in which a portion of the shift elements are implemented as form-locking shift elements.
Due to the design of a portion of the shift elements of a transmission as form-locking shift elements, the power loss due to the drag torque of disengaged shift elements is reduced, and the transfer capability is increased with respect to the shift elements, wherein the mechanical overall gear ratio spread remains the same.
The mechanical overall gear ratio spread of a transmission is a key control variable for operating the upstream internal combustion engine at an optimal operating point, whereby the fuel consumption can be reduced. Further, the mechanical overall gear ratio spread of a transmission is an important parameter for attaining a specific driving performance in special applications.
Further developments of existing transmissions known from the prior art, implemented to be shiftable under load, result in a slight increase of the overall gear ratio spread and are disadvantageously complex and expensive.
Further from the prior art, it is known to combine automatic transmissions with additional automatically shiftable distributor transmissions for representing a group shift, which are integrated in the drive strategy, thereby increasing the mechanical overall gear ratio spread.
However, this design has the disadvantage that a two-stage distributor transmission is necessary for representing a group shift, which results in high manufacturing and assembly costs and large construction space needs. The distributor transmission, as a rule, is implemented as a transmission in countershaft design. Additionally, the distributor transmission, with respect to the attainable transmission ratio, the design of the form-locking shift elements and the synchronization measures, must be adapted to the upstream transmission and the internal combustion engine. In addition, the group shifts, disadvantageously, cannot be shifted under load.
Transmissions of road vehicles are used to some extent with special applications in the non-road range, such as rail cars or special-purpose vehicles (rail cars, motor boats, special-purpose vehicles in the high off-road range) in order to minimize costs. Bus transmissions are used particularly for applications in the high power range with rail car and special-purpose vehicle applications. Because these transmissions are adapted specifically to the requirements of buses, for the most part they have only one reverse gear, and for special applications, the transmission ratio spread is too low. For rail cars and for some special applications however, the same number of forward and reverse gears are necessary. Reversing transmissions can be used in this context.
The document EP 0 965 773 A1 from the applicant describes a reversing transmission in countershaft design. Such reversing transmissions are heavy, require a large construction space and result in high manufacturing and assembly costs. A further reversing transmission is disclosed in the document DE 10 2010 039 862 A1 from the applicant; the disclosed reversing transmission is designed as a separate assembly, and is disposed in the power flow direction in tractive mode behind a transmission.